Known as a fuel injection pump for a diesel engine is the one having a configuration shown in FIG. 4. The fuel injection pump shown in FIG. 4 is a so-called distributor type fuel injection pump based on an electronic control system. In FIG. 4, 10 refers to a pump main body, and a vane type feed pump 11 is disposed within the pump main body 10. Here, as for the feed pump 11, together with its original side view, a front view with an angle of representation changed by 90.degree. is also shown.
The feed pump 11 is rotated by a drive shaft 12 which is actuated as the engine rotates, thereby forcibly feeding fuel from a fuel tank. The fuel emitted from the feed pump 11 is transmitted to a pump chamber 13 within the pump main body 10, and then is supplied therefrom to a fuel forced feed plunger 15 through a passage 14. Inserted into the passage 14 is a fuel cutting magnet valve 16.
The plunger 15 supplies, by way of a communicating hole 17A formed therein, fuel from the passage 14 to a delivery valve 19 through a passage 18, while moving back and forth within a plunger chamber 17 formed in the pump main body 10. Such back and forth movement of the plunger 15 is effected by a cam disc 20 connected to one end of the plunger 15.
Namely, the plunger 15 and the cam disc 20 are rotated by the drive shaft 12 in response to the engine rotation. Also, the cam disc20 is urged by a spring 21 via the plunger 15, thereby abutting to a roller 23 axially supported by a roller holder 22. Here, the roller holder 22 does not move in the axial direction of the drive shaft 12 and usually (except for the time when rotational phase is being adjusted, which will be explained later) does not rotate around the axis of the drive shaft 12. Consequently, the cam disc 20 moves in the axial direction while being shoved by the roller 23 according to its cam profile. Thus, the plunger 15 moves back and forth, thereby supplying fuel at a desired timing.
Here, each cylinder is provided with the passage 18 and the delivery valve 19. For example, in the case of a four-cylinder engine, four pieces each of passages 18 and delivery valves 19 are provided.
As shown in FIGS. 5, the roller holder 22 is provided with a plurality of rollers 23 (which are four here), and the cam disc 20 has a cam profile corresponding thereto. Consequently, as the cam disc 20 makes one revolution, the plunger 15 is driven four times, whereby, for example, fuel is successively supplied to four cylinders respectively in response to these four driving operations of the plunger 15.
Here, provided for fuel injection amount control are a control sleeve 24, which moves back and forth on the outer periphery of the plunger 15 so as to adjust the forced feed stroke of the plunger 15, and a governor (electric governor here) 25 for controlling the control sleeve 24.
Further, in FIG. 4, 26 is a regulator valve, 27 is a sensing gear plate for detecting the rotational speed of the drive shaft 12, 28 is a fuel temperature sensor, and 29 is an overflow valve, provided with a check valve, for returning excess fuel within the pump chamber 13 to the fuel tank.
In order to control the fuel injection timing, such a fuel injection pump is provided with a timer 30. The timer 30 is equipped with a timer piston 31 for changing the position of the roller 23 in its rotating direction. Here, for convenience, the timer piston 31 is also depicted by a front view whose angle of representation is changed by 90.degree..
As shown in FIGS. 4, 5(A), and 5(B), the timer piston 31 minutely rotates the roller holder 22 via a piston pin 33 while moving back and forth within a cylinder 32 formed in the pump main body 10.
Namely, the timer piston 31 has an intermediate portion to which the piston pin 33 is connected, one end provided with a first pressure chamber 34 into which the fuel pressure within the pump chamber 13 is introduced, and the other end provided with a second pressure chamber 35 into which the intake-side fuel pressure (fuel pressure upstream the feed pump 11) is introduced.
Also, the timer piston 31 is provided with a passage 36 through which the pump chamber 13 and the first pressure chamber 34 communicate with each other, and the passage 36 is formed with an orifice 37. Further, disposed within the second pressure chamber 35 is a timer spring 38 for urging the timer piston 31 toward the one end (toward the first pressure chamber 34).
Hence, the position of the timer piston 31 is determined according to the balance among the fuel pressure within the first pressure chamber 34, the fuel pressure within the second pressure chamber 35, and the urging force of the timer spring 38. For example, when the fuel pressure within the first pressure chamber 34 becomes higher than that in the state shown in FIG. 5(A), the timer piston 31 moves to the left in the drawing as shown in FIG. 5(B), whereby the fuel injection timing is adjusted to the advancing side. When the fuel pressure within the first pressure chamber 34 becomes low, by contrast, the timer piston 31 moves to the right in the drawing, whereby the fuel injection timing is adjusted to the retarding side.
For example, when the rotational speed of the engine becomes high, the output pressure from the feed pump 11 increases, whereby the fuel pressure within the pump chamber 13 also increases, thus yielding a high pressure within the first pressure chamber 34. Consequently, the timer piston 31 moves to the left in the drawing, so that the fuel injection timing is adjusted to the advancing side.
Further, as shown in FIG. 4, disposed in the case of this pump is a timing control valve (TCV) 39 which can adjust the pressure balance between the first pressure chamber 34 side and the second pressure chamber 35 side, whereby the fuel injection timing can be adjusted on the basis of various parameters.
Namely, the timing control valve 39 is a solenoid valve of electronic control type, whose opening (valve opening time per unit time) is adjusted by duty control. Accordingly, as the timing control valve 39 is duty-controlled, the pressure difference between the pressure on the first pressure chamber 34 side and the pressure on the second pressure chamber 35 side is appropriately adjusted (to the reducing side here) according to the opening of the valve 39, thus regulating the position of the timer piston 31, whereby the fuel injection timing is adjusted.
Such driving of the timing control valve 39 is effected by a non-depicted timing control valve driver (TCV driver), whose operation is controlled by a non-depicted controller according to a target fuel injection amount Q and an engine rotational speed (i.e., number of engine revolution per unit time, which will be hereinafter referred to as engine speed) Ne.
In the case where the timing control valve 39 is duty-controlled, such control is effected while driving current pulses are emitted at a predetermined frequency. When this driving frequency approaches an integral multiple of the engine speed, the fuel injection timing may fluctuate. This phenomenon is known as "fluctuation," which is supposed to be a phenomenon of so-called "beats" caused by interference of two different frequencies close to each other.
For example, FIG. 6 shows experimental data concerning this fluctuation phenomenon, in which results of an experiment under a condition where the engine speed Ne is in the vicinity of 1,800 rpm and the driving frequency of the timing control valve 39 is 60 Hz. In FIG.6, the abscissa and ordinate respectively indicate time and position of the timer piston 31 (TPS), whereas curve S indicates a fluctuation characteristic of the TPS, from which it can be seen that the TPS is vibrating with a relatively long period (about 1 second).
It is supposed that, while the driving frequency of the timing control valve 39, i.e., 60 Hz, becomes just twice that of the engine speed Ne when the latter is exactly 1,800 rpm (=30 Hz), such a phenomenon as "beats" occurs due to the fact that the engine speed Ne is close to but not exactly 1,800 rpm.
Thus simulated is a case where the engine speed Ne is 1,700 rpm (=29.5 Hz) while the driving frequency of the timing control valve 39 is set to 60 Hz, which yields, as shown in FIG. 7, a characteristic substantially the same as the experimental results shown in FIG. 6.
When the orifice 37 constantly acts, and the opening of the timing control valve 39 is kept constant, the inflow/outflow of fuel in the individual pressure chambers 34 and 35, i.e., the position of the timer piston 31, is ruled by the pressure difference between both ends of the timer piston 31. Accordingly, changes in pressure difference are supposed to cause the fluctuation (i.e., displacement of the timer piston 31).
Further, presumed to be factors for changes in pressure difference are pressure changes in the pump chamber 13 and pressure changes in the cylinder 32 of the timer piston 31.
The pressure may change in the pump chamber 13 due to changes in the discharge pressure of the feed pump 11, spills of the fuel forcibly fed by the plunger 15, and the like. Also, the pressure may change in the timer piston cylinder 32 due to the fact that the reaction force generated when the cam disc 20 runs over the roller 23 is transmitted through the piston pin 33, due to the resonance generated between the mass of the timer piston 31 and the elastic property of the timer spring 38, and the like.
Among them, particularly influential is the reaction force generated by the cam disc when it runs over. Simply put, this reaction force is twice as influential as the change in pressure of the pump chamber 13, since the inflow pressure changes at both ends of the timer piston 31 (i.e., in both pressure chambers 34 and 35) when the cam disc runs over. Also, it can be considered most influential since the forced pressure acting on the timer piston 31 itself fluctuates greatly. In the case of a four-cylinder engine, the reaction force generated by the cam disc when it runs over has a frequency twice that of the engine speed Ne and a substantially constant amplitude.
In order to eliminate such a fluctuation phenomenon, the following means can be considered:
(1) Namely, as indicated by lines L1, L2, and L3 in FIG. 8, the driving frequency of the timing control valve 39 is completely synchronized with a resonance point with respect to the engine rotation, i.e., the engine speed Ne, the level twice as high as the engine speed Ne (=2Ne), or the level four times as high as the engine speed Ne (=4Ne). This technique is disclosed, for example, in Japanese Patent Publication No. SHO 63-8298.
(2) The driving frequency of the timing control valve 39 is prevented from approaching the resonance point with respect to the engine rotation, i.e., the engine speed Ne, 2Ne, or 4Ne, while being changed with respect to the engine speed Ne like sawteeth as indicated by dotted line L4 in FIG. 8. This technique is disclosed, for example, in Japanese Patent Publication No. HEI 1-19059.
These means are based on a characteristic in which the greater is the frequency difference between the timing control valve and the engine, the smaller becomes the fluctuation (piston amplitude). In FIG. 9, a characteristic indicated by Ne refers to a case in the vicinity of the 0.5-order resonance point, i.e., where the driving frequency of the timing control valve is set near the engine speed Ne; that indicated by 2*Ne refers to a case in the vicinity of the first-order resonance point, i.e., where the driving frequency of the timing control valve is set near the level twice as high as that of the engine speed Ne; and that indicated by 4*Ne refers to a case in the vicinity of the second-order resonance point, i.e., where the driving frequency of the timing control valve is set near the level four times as high as that of the engine speed Ne.
Regarding the solving means (1), the above-mentioned Japanese Patent Publication No. SHO 63-8298 relates to a technique in which the operating frequency of an opening/closing valve (timing control valve) is controlled, according to the fuel feed pressure, so as to become an integral multiple (e.g., twice that) of the engine speed, thereby synchronizing the operating period of the opening/closing valve with the pressure-changing period in the fuel feed pressure, thus decreasing the fluctuation in fuel feedpressure and accurately controlling the fuel injection timing.
In this technique, however, since the fuel feed pressure is detected so as control the operation of the opening/closing valve on the basis of thus detected pressure, there are disadvantages as follows:
First, as to the above-mentioned fluctuation phenomenon, the control based on the fuel feed pressure is not always appropriate since the fuel feed pressure is relatively less influential.
Also, the timing for starting the driving of the opening/closing valve changes every time the fuel injection timing changes, thus complicating the setting for the timing for terminating the driving of the opening/closing valve (i.e., duty width). When the driving terminating timing for the opening/closing valve is to be synchronized with the rotational phase, on the other hand, the driving starting timing for the opening/closing valve cannot be set correctly, whereby the opening control of the opening/closing valve (i.e., duty control) cannot be achieved.
Further, there is needed a contrivance for synchronizing the signal based on the detected fuel feed pressure with the crank angle of the engine. It is due to the fact that the rise of fuel feed pressure synchronizes with the injection timing but does not relate to the crank angle, and that the fall of fuel feed pressure depends on the amount of injection but does not relate to the crank angle either.
In the technique of (2), while a plurality of driving frequencies are selected according to the engine speed (rotational speed), the effect of suppressing the fluctuation phenomenon cannot securely be obtained unless points for switching the driving frequencies are appropriately set.
For example, in a switching point P1 within the first-order resonance region in the vicinity of the line L2 in FIG. 8, the driving period of the timing control valve and the engine speed approach the first-order resonance region, whereby the above-mentioned fluctuation phenomenon tends to occur greatly. It is due to the fact that the piston amplitude caused by fluctuation becomes greater in the first-order resonance region.
Namely, as indicated by the characteristic line 2*Ne in FIG. 9, the piston amplitude in the vicinity of the first-order resonance is substantially twice as large as that in the vicinity of the 0.5-order resonance (characteristic line Ne) and nearly four times as large as that in the vicinity of the second-order resonance (characteristic line 4*Ne).
Thus, the greater the piston amplitude is, the lower becomes the accuracy in controlling the fuel injection timing. Accordingly, the fluctuation phenomenon in the first-order resonance region should be eliminated in particular. The prior art has conceived no particular means against such fluctuation in the first-order resonance region, thus failing to securely attain the effect of suppressing the fluctuation phenomenon.
Though Japanese Patent Application Laid-Open (Kokai) Nos. HEI 1-300037 and 4-347346, and Japanese Patent Publication No. HEI 3-25626 each disclose a technique concerning the fuel injection timing control for engines such as diesel engine, they fail to specifically take account of the fluctuation phenomenon such as that mentioned above and cannot appropriately eliminate such fluctuation phenomenon.
In view of the problems mentioned above, it is an object of the present invention to provide, in an apparatus which can control a fuel injection timing by adjusting the position of a timer piston via a solenoid valve, an engine fuel injection timing control apparatus which can securely attain an effect of suppressing the fluctuation phenomenon, thus allowing the fuel injection timing to be controlled appropriately.